The internal combustion engine of a portable handheld work apparatus such as a motor-driven chain saw, a cutoff machine, suction/blower apparatus or the like generates oscillations in its rpm range. These oscillations are, for example, noticeable as vibrations in a handle for guiding the work apparatus. A further component of the oscillatory excitation is regularly generated by the tool which is driven by the engine. The tool can, for example, be a saw chain, a cutoff disc, a cutting knife or the like. The oscillations generated thereby in the handle can lead to a premature tiring of the operator.
The motor unit of a portable handheld work apparatus and a unit, which is isolated relative to the work apparatus with respect to vibration (for example, with a handle), have comparatively small masses in order to make it possible to easily carry the apparatus and to easily guide the same with the hand during operation. The small light drive motor is operated at high engine speed (rpm) to generate an adequate drive power. Because of the low rotational inertial forces of the small drive motor, its idle rpm is also comparatively high. In total, an excitation frequency spectrum results which is high. The excitation force amplitudes are likewise high because of the uneven engine running of the mostly one-cylinder drive motor referred to the weight forces of the components to be insulated.
There are many embodiments of work apparatus known wherein, for example, a handle is fixed to the motor unit of the work apparatus via an antivibration device. The antivibration device is intended to provide vibration insulation of the handle from the motor unit. One such antivibration device includes a rubber vibration damper with combined elastic and damping characteristics. A decoupling of vibration can be adjusted via a targeted dimensioning of the elastic characteristics. A portion of the vibration amplitudes, which are transmitted nonetheless to the handle element, can be damped by the material characteristics of the rubber.
The essentially non-linear material characteristics of the rubber can be disadvantageous in this context. For example, the stiffness of a rubber element increases with larger deflections and is essentially caused by its significant transverse expansion. As a consequence, the resonance frequency of the vibrating system made up of the motor unit, the handle, the intermediately connected antivibration element or device changes in dependence upon the preload and the vibration amplitude. An adaptation of the resonance frequency to the operating frequency range of the work apparatus is therefore difficult. An operation of an antivibration element of this kind in a quasi-linear range is only possible for a correspondingly large configuration of the antivibration element for which sufficient mounting space is not always available. High operating loads or tight spatial conditions require the arrangement of a vibration damper, for example, in a sleeve, which prevents the transverse expansion of the damper material. Blocking the transverse expansion leads, with rubber, to a considerable stiffening, which makes an adaptation to the excitation frequencies to be dampened difficult.
A further disadvantage of rubber as a material for a damping element lies in its frequency-dependent stiffness. At high frequencies, the elasticity module of the rubber material increases. An increased inherent frequency can adjust which leads to resonances at the comparatively high excitation frequencies. Under unfavorable conditions, the resonance frequency can even migrate with the excitation frequency. Unwanted resonances over a wide excitation frequency range can occur. For an adequate vibration decoupling at high excitation frequencies, a very soft dimensioning of the antivibration element is required, which, under some circumstances, can lead to an excessively soft connection of the handle element to the motor unit. A clean guidance of the work apparatus is therefore hindered.
Further disadvantages can occur because of the stiffening of the rubber material at low temperatures or because of deterioration. A constructively pregiven vibration decoupling can then, under some circumstances, no longer be achieved in practice.
In alternate embodiments, antivibration elements having steel springs are known whose spring characteristics are essentially constant or linear. However, the low material damping of the steel is here disadvantageous and can lead to unwanted resonances. An antivibration element having a vibration damper of steel is furthermore sensitive with respect to material fatigue.
U.S. Pat. No. 6,471,179 discloses to use isolation elements made of a polyurethane foam for a vibration-damping attachment of a vehicle chassis to a vehicle body. A corresponding vehicle chassis can, especially in the loaded state, exhibit a considerable inherent weight which charges the isolation elements with a high static base load. The isolation elements must exhibit a correspondingly high carrying capacity.
The vibration-capable system, which is formed in this way, is subjected essentially to low-frequency excitations which can arise when driving on an uneven roadway. The low frequency excitations caused by the roadway unevenness form an excitation spectrum which does not change with the loading state of the vehicle's body.
PU-foam isolation elements first become softer with increasing static loading until a stiffness minimum is reached. With a further increase of the static load, an increasing hardening then develops.
There is also a dynamic hardening for PU-foam isolation elements to be observed. For a dynamic, vibrating load, the stiffness increases with the load frequency. The increase of the dynamic stiffening is most pronounced at low frequencies.
For a correct static and dynamic dimensioning of the suspension of the vehicle body by means of PU-foam isolation elements, the PU-foam, in the unloaded state of the vehicle body, is so greatly statically pretensioned by its own weight that approximately the static stiffness minimum is reached. In combination with the empty mass of the vehicle body, a constructive pregivable inherent frequency of the vibration-capable system adjusts.
In the loaded state, the mass of the vehicle body and the static loading of the isolation elements are increased. The stiffness of the PU-foam increases. In combination with the increased mass of the vehicle body, the natural frequency can at least be held approximately constant independently of the static loading state. The natural frequency of the vibration system can be determined under static considerations. However, the PU-foam damper performs in this range like a conventional rubber damper whose stiffness likewise increases with the static load. With respect to the dynamic matching, no special advantages can be achieved with the use of a PU-foam damper compared to a rubber damper.
The high static loading because of the weight force of the vehicle body requires a sturdy dimensioning of the elastic dampers. Compared to the low-frequency excitation when traveling over uneven ground, a high natural frequency of the vibrating vehicle body occurs. For the frequency range of a few Hz in question, the especially pronounced dynamic stiffness of the foam material here contributes to the increase of the natural frequency. The vehicle body is essentially subjected to an undercritical excitation.
Via the chassis of the vehicle, higher excitation frequencies can, however, also be transmitted to the vehicle body and unwanted resonances can occur.
This effect can be amplified when the PU-foam dampers are designed for low loads. For the unloaded vehicle body, the stiffness minimum of the foam material is not yet reached. With increasing loading of the vehicle body, the stiffness of the PU-foam becomes less. The natural frequency of the system drops. Low frequency excitations during travel over an uneven roadway can also lead to resonances. In this case, the use of the PU-foam dampers is even disadvantageous compared to rubber dampers.